Compressors

ABSTRACT

Disclosed compressor embodiments can prevent or substantially reduce the inefficient upwardly spiraling heat buildup found in typical compressors by significantly increasing the surface area of mating piston and chamber walls, using the meshing and undulating geometry of the piston/chamber walls to create increased air turbulence in the chamber and generate more air flow against the cooling walls, promoting turbulence even more by varying the axial path of the piston head so that it is not simply linear, using fluid to cool the non-working side of the piston, and/or providing additional cooling features (such as fins) on the non-working side of the piston to provide still further cooling means for the piston head. Analogous principles can also be included in wobble-plate compressors, scroll compressors, blowers, and other compressor types.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of U.S. Provisional Patent Application No. 61/961,553, filed Oct. 18, 2013, and U.S. Provisional Patent Application No. 61/962,523, filed Nov. 12, 2013, both of which are incorporated by reference herein in their entirety.

FIELD

This disclosure is related to air compressors, pneumatic motors, and related devices, and particularly to near-isothermal embodiments and/or those having a high degree of efficiency.

BACKGROUND

Compressors for compressing fluids, such as air, are inefficient. The amount of energy consumed by such compressors greatly exceeds the theoretical amount of work done merely to compress the fluid. The remainder of the consumed energy ends up as unwanted waste heat.

The true inefficiency of practical compressors appears not to be widely known. Discussions of compressor inefficiency in the literature tend to compare isothermal compression versus adiabatic compression, with isothermal compression being theoretically the most efficient compression. Adiabatic compression is sometimes thought to be the worst case possible. The calculations are usually done assuming reversible processes. A reversible adiabatic 10:1 volume reduction for air requires only 1.63 times as much as or 63% more work than a reversible isothermal 10:1 volume reduction for air.

However, actual compressors, working against their stated maximum working pressure and fully warmed up to steady-state operation, are far more inefficient than reversible adiabatic theory suggests. A truer indication of how compressors tend to operate appeared in a recent advertisement in an engineering journal. Kaeser Compressors, Inc., a major manufacturer of industrial compressors and blowers, noted, “It is a fact: Compressed air is inherently inefficient. It takes 8 kW of electricity to deliver 1 kW of power in compressed air—and almost all of the remaining 7 kW is lost as heat.” By this one measure, the energy penalty in industrial compressors is about 700%. Some of this loss, about 15% of the total 8 kW, is due to friction, and not to the inefficiency of the compression process itself. That leaves a penalty of about 5.8 kW, or 73% of the energy used in the best compressors, being lost as waste heat. This is not how compressors tend to be described and discussed in the literature.

The wastefulness of real-world compressors has long been lamented, and many have opined that if only an isothermal compressor were possible, much energy could be saved. But, an isothermal compressor is widely believed to be impossible. So much so that the concept of an isothermal compressor has been described as something of a “Holy Grail.” Conventional wisdom suggests that development of an isothermal compressor is akin to trying to develop a perpetual motion machine, or turn lead into gold.

Contrary to conventional wisdom, thermodynamic theory indicates that an isothermal compressor is possible, but only if the compressor is operated infinitely slowly, so that the amount of heat that must flow out of the gas during isothermal compression can flow out reversibly. If any compressor could operate infinitely slowly, it would remain in thermal equilibrium with its surroundings, and its operation would be truly isothermal. Thus, almost any common piston-type compressor, for example, could be run isothermally at near zero rotational speed (assuming perfect sealing, so that the developed pressure is not lost). So, an isothermal compressor is not impossible. It is just not possible in any practical, useful sense. A compressor that requires an infinite amount of time per compression cycle is not practical; in fact it is of no value at all.

If means can be found to keep a compressor and the compressing gas in near-thermal equilibrium with the surroundings, but at operational speeds that are useful, something like isothermal compression can be achieved. A problem which conventional wisdom fails to recognize is that the heat developed by the compression process must be drawn out of the gas quickly enough, and during the compression if near-isothermal compression is to be achieved. Thus, developing a practical near-isothermal compressor is not just a thermodynamics problem; it is also kinetics problem.

It is quite a kinetics problem. Reciprocating (piston-type) compressors, for example, tend to run in the 800 to 1,000 rpm range. At 900 rpm, the compression stroke lasts only 33 milliseconds (0.033 seconds). If the heat of compression is going to be extracted from the gas during the compression, in emulation of isothermal operation, the heat has to be extracted in a fraction of this short time span. Perhaps something like 20 milliseconds is available during each compression stroke to remove the heat, or some large fraction of it.

Another common misconception is made by those who assume that an isothermal compressor, and nothing less, is what is needed. It is not. Almost all of the value that would be obtained from a fully isothermal compressor can be had from a compressor that is only close to isothermal. In fact, it is not even necessary to get very close. Any approximation of isothermal operation would reap large benefits.

The reason for this is the compounding nature of the temperature rise that occurs during adiabatic compression. The initial part of the compression goes reasonably well, with little increase in temperature. But, the temperature rise feeds on itself. The higher the temperature, the faster the temperature rises. The losses begin to compound. In the end, even a modest compression of a fluid, say from 15 psi to 150 psi, results in a temperature rise of hundreds of degrees.

As a specific example, compressing a quantity of a gas from 70° F. and 15 psi to just over three times the pressure, 50 psi, results in a temperature rise of 335 degrees Fahrenheit, to 405° F. But, triple the pressure again, to 150 psi, and the temperature rises another 477 degrees Fahrenheit, to 882° F. Triple the pressure again, to 450 psi, and the temperature rises another 739 degrees Fahrenheit, to 1,621° F. An aluminum compressor piston would melt. Yet, these are fairly modest pressures.

One conventional method used to interfere with the temperature rise during compression is water injection. Water injection is currently used in some commercially available screw compressors, and experiments have been done with water injection in reciprocating (piston-type) compressors. Some energy savings have been achieved. However, there are technical challenges with water injection. One disadvantage of water injection is that it adds significant water to the compressed gas. For most uses, the water has to be removed, and dehumidification requires a refrigerator-type drier that consumes significant energy (using, incidentally, a compressor to achieve the cooling).

One conventional attempt to interrupt the temperature rise is to use multi-stage compressor trains, with inter-stage cooling of the gas. Depending on the final pressure required, two-stage, or even three-stage compressors can be more economical than single-stage compressors, in spite of the additional complexity, friction, maintenance, and capital investment. (Pressures above 150 psi are generally not attempted with single-stage compressors, due to excessively high operating temperatures.)

One reason why little effort (other than water injection) has been made to interfere with the temperature rise in compressors is that as the gas is compressed in the compression chamber (a cylinder in piston-type compressors), it heats up essentially homogeneously. Except for the gas that is immediately close to the walls of the compression chamber (stator, rotor, scroll, cylinder wall, piston, head, etc.), all of the gas heats throughout at the same rate; and, there seems to be no means possible to stop the heating (other than water injection).

The problem is that only the gas within a few mean free paths of the walls of a compressor can lose some heat to the walls and thus avoid some of the temperature rise that would otherwise occur. The mean free path of air molecules, for example, at one atmosphere pressure is on the order of 70 nM or about two and a half millionths of an inch. At higher pressures, it is proportionally less. Only the gas that is within a few microns of the walls of a compressor is in thermal contact with the walls. The remainder of the gas is thermally insulated from the walls by the gas next to the walls. The great majority of the gas in a compressor is well insulated from the (potentially cooler) walls. Stagnant air, it turns out, is a good thermal insulator. This is why fiberglass insulation works so well in homes. The fiberglass itself provides almost no insulating value; it is the layer of stagnant air it creates that is the actual insulator.

This helps explain why conventional attempts to cool compressors, such as the use of water jackets, cooling fins and cooling fans used to cool the outer wall of the compression chamber, are sufficient to allow compressors to function adequately, sometimes continuously, but do little to address the fundamental efficiency problem of compressors.

More specifically, because most of the gas is well insulated from the cooler chamber wall of a well cooled compressor, the gas still begins to rise in temperature as the compression begins, especially in the center of the gas. This causes excess work to be done for further compression. Which causes the temperature to rise further and faster, in a very quick upward spiral, as the gas is suddenly compressed. The result is a highly inefficient operation.

In conventional compressors, the compressed gas may lose some of its heat to the cooler walls as it flows turbulently out of the compression chamber.

However, post-compression cooling of the gas, once the excess work is done and the game is lost, is not a good solution. And such post-compression cooling results in a pressure drop which is counter-productive. What is needed is a way to avoid the excess work, to stop the upward temperature spiral as it is happening, or better yet, to trip it up, so that it does not get started in the first place. Contrary to conventional approaches, the solution is to interfere with the compression heating of a gas in a compressor, and thereby at least partially avoid it, rather than dealing with its consequences after the fact.

Such a solution is not hopeless, as commonly believed, because the amount of heat that must be removed from a gas as it is isothermally compressed is quite small. Real-world compressors generate so much heat, and that heat is such a monumental problem, that it is easy to believe, incorrectly, that a large amount of heat must be extracted from the gas as it is compressed to achieve isothermal compression. This is far from true. The actual amount of heat lost by a gas in a theoretical, infinitely slow isothermal compression is actually quite small. All the other heat that we associate with conventional compressors is excess heat that need not be generated at all. The excess heat largely is a result of the upward spiral of temperature rise, with each succeeding increment of compression being harder than the last due to the rising temperature and requiring even more work. It is a perfect ratcheting upward of something that should be avoided altogether, if possible.

As an example, it is easy to calculate for a very common, rather small compressor that the theoretical, isothermal-equivalent work done by the compressor is a fraction of the total amount of shaft horsepower the compressor consumes from an electric motor in actual operation. Assuming the following technical specifications: Bore—2.375″, Stroke—2″, Rotational Speed 919 RPM, Maximum Operating Pressure—125 psi, Flow Rate at Maximum Operating Pressure—1.54 scfm (Factory claim=33% volumetric efficiency), Power Requirement—1 Horsepower, the isothermal-equivalent work being done is approximately ⅓ of a horsepower. And, this is the full amount of heat that needs to be extracted from the gas, if the compression can be kept isothermal throughout the stroke. All other heat generated by this compressor is unnecessary. ⅓ of a horsepower is only 250 Watts. It is quite easy to deal with 250 Watts of heat: that amount of heat put into one liter of water for a full minute raises the temperature of the water by only 3.6 degrees Centigrade or 6.5 degrees Fahrenheit.

SUMMARY

One aspect of the disclosed technology is to significantly increase the inner surface area of the piston head and compression chamber to allow for much greater cooling of the air during the compression cycle, thereby disrupting the rapidly compounding heat buildup with each incremental increase in compression. In some embodiments, the piston head and chamber are provided with opposing interleaving or nesting projections that provide the piston head and chamber walls with much greater surface area (compared to a typical cylindrical piston design) to cool the air much more effectively as it compresses. The moving piston can still move axially within the compression chamber, allowing the mating piston and chamber walls to compress the air as much as is needed and yet also act as a more effective cooling means. As a another benefit, the mating projections/protuberances on the piston head and chamber wall surfaces also create greater air turbulence during the compression stroke, thereby causing a larger volume of air to pass closely by the wall surfaces and be cooled.

Another aspect of the disclosed technology is to laterally or angularly shift or rotate slightly the piston during the compression stroke such that the piston path is no longer perfectly axial. This more complex movement differs from the traditional linear axial stroke of conventional compressor pistons and creates more air turbulence within the chamber, thereby promoting even more cooling as a larger volume of air passes in close proximity to the cooling wall surfaces of the piston and chamber.

Still another aspect of the disclosed technology is the use of a fluid (oil or water, for example) to internally cool the piston head and ultimately the compressed air in the chamber. Some embodiments include a cooling oil spray on the internal non-working side of the piston. Some embodiment include additional internal (i.e., non-working side) cooling features, such as cooling fins, projections, protuberances, etc. For example, the cooling fluid can be sprayed against the internal cooling fins, projections, etc. on the non-working side of the piston to provide even greater cooling of the piston and ultimately the compressed air.

As a whole, disclosed compressor embodiments described herein can prevent or substantially reduce the inefficient upwardly spiraling heat buildup found in typical compressors by significantly increasing the surface area of mating piston and chamber walls, using the meshing and undulating geometry of the piston/chamber walls to create increased air turbulence in the chamber and generate more air flow against the cooling walls, promoting turbulence even more by varying the axial path of the piston head so that it is not simply linear, using fluid to cool the internal non-working side of the piston, and/or providing additional cooling features (such as fins) internal to the piston to provide still further cooling means for the piston head.

The foregoing and other objects, features, and advantages of the disclosed technology will become more apparent from the following detailed description, which proceeds with reference to the accompanying figures.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a perspective view of an exemplary compression chamber geometry.

FIG. 1A is a partially cross-sectional side elevation view of an exemplary compressor.

FIG. 1B is an enlarged view of a portion of FIG. 1A.

FIG. 2A is a partially cross-sectional side elevation view of another exemplary compressor.

FIG. 2B is an enlarged view of a portion of FIG. 2A.

FIG. 3 is a perspective view of a portion of an exemplary compressor.

FIG. 4 is an exploded view of the compressor of FIG. 3.

FIG. 5 is a side view a piston and a housing head of the compressor of FIG. 3, showing opposing projections nested together.

FIG. 6 is a perspective view of the housing head of the compressor of FIG. 3.

FIG. 7 is a side view of the housing head of the compressor of FIG. 3.

FIG. 8 is a cross-section bottom view of the housing head taken along section line 8-8 of FIG. 7.

FIG. 9 is a perspective view of the piston of the compressor of FIG. 3.

FIG. 10 is a bottom view of the piston of the compressor of FIG. 3.

FIGS. 11A-11C illustrate isolated non-axial relative motion between opposing surfaces in a compression chamber and resultant gas flow therebetween.

FIG. 12 illustrates a combination of axial compression motion and non-axial oscillation motion between opposing projections in a compression chamber.

FIG. 13 is a partially cross-sectional view of an exemplary wobble plate-type compressor.

FIG. 14 is an enlarged view of cooling projections in the compressor of FIG. 13.

FIG. 15A shows a port of the compressor of FIG. 13 in an intake configuration.

FIG. 15B shows the port of the compressor of FIG. 13 in an exhaust configuration.

FIG. 16 is a schematic illustration of an exemplary scroll-type compressor.

FIGS. 17A and 17B are cross-sectional views of an exemplary scroll-type compressor represented by FIG. 16, taken along section lines 17A-17A and 17B-17B, respectively, of FIG. 16.

FIGS. 18A and 18B are cross-sectional views of another exemplary scroll-type compressor represented by FIG. 16, taken along section lines 18A-18A and 18B-18B, respectively, of FIG. 16.

FIG. 19 is a partially cross-sectional view of an exemplary blower-type compressor.

FIG. 20 is a cross-sectional view of the blower-type compressor of FIG. 19, taken along section lines 20-20 in FIG. 19.

FIG. 21 is an enlarged view of a portion of FIG. 19, showing an interface between nesting projections of the housing and the rotor.

DETAILED DESCRIPTION

For purposes of this description, certain aspects, advantages, and novel features of the embodiments of this disclosure are described herein. The disclosed methods, apparatuses, and systems should not be construed as limiting in any way. Instead, the present disclosure is directed toward all novel and nonobvious features and aspects of the various disclosed embodiments, alone and in various combinations and sub-combinations with one another. The methods, apparatuses, and systems are not limited to any specific aspect or feature or combination thereof, nor do the disclosed embodiments require that any one or more specific advantages be present or problems be solved.

Although the operations of some of the disclosed methods are described in a particular, sequential order for convenient presentation, it should be understood that this manner of description encompasses rearrangement, unless a particular ordering is required by specific language. For example, operations described sequentially may in some cases be rearranged or performed concurrently. Moreover, for the sake of simplicity, the attached figures may not show the various ways in which the disclosed methods and devices can be used in conjunction with other methods and devices.

As used herein, the terms “a”, “an” and “at least one” encompass one or more of the specified element. That is, if two of a particular element are present, one of these elements is also present and thus “an” element is present. The terms “a plurality of” and “plural” mean two or more of the specified element.

As used herein, the term “and/or” used between the last two of a list of elements means any one or more of the listed elements. For example, the phrase “A, B, and/or C” means “A,” “B,” “C,” “A and B,” “A and C,” “B and C” or “A, B and C.”

As used herein, the term “coupled” generally means mechanically, chemically, or otherwise physically coupled or linked and does not exclude the presence of intermediate elements between the coupled or associated items absent specific contrary language.

As used herein, the term “compressor” means a device that use mechanical work to compresses a gas, such as air, and the term excludes devices, such as internal combustion engines, that include or operate using combustion or other chemical release of energy within a chamber.

Compression Chamber Geometry

The following description recognized that, in order to minimize the compression heating that occurs in a compressor, and thereby increase the efficiency of the compressor, it is very beneficial to extract heat from throughout the gas in the compression chamber directly into surfaces having high thermal conductivity. In a traditional cylinder shaped compression chamber, a relatively low percentage of the gas in the chamber is adjacent to the walls of the chamber, and the large percentage of the gas in the middle of the chamber is only cooled via relatively inefficient conduction of heat from the central volume of the gas to the outer walls of the chamber. This leads to poor cooling of the gas in the middle of the chamber. To more efficiently extract heat from the entire gas volume in the chamber, additional cooling surfaces can be placed inside the compression chamber such that the gas near the center of the chamber is closer to the nearest surface and can transfer heat to the closer surfaces faster.

A conventional cylindrical cylinder shape with flat upper and lower walls makes sense from the perspective of maximizing the total volume of the cylinder and keeping it free of complicating features that take up space, space that can otherwise be filled with more gas. However, the principles and designs disclosed herein recognize that surface features, such as projections or protuberances, that extend from the working face of the piston or from the opposing face of the head or from the lateral walls of the compression chamber do not necessarily result in less usable room. In fact, in a piston-type compressor, the working volume of the compression chamber is determined by the cross-sectional area of the chamber and the stroke length of the piston (referred to as the swept volume). The piston and the head can have any shapes, without subtracting from or adding to the working volume or clearance volume of the cylinder, as long as a sufficient cross-sectional area of the chamber and a sufficient stroke length is provided. In addition, if an irregular shape on the working face of the piston nests matingly with a corresponding opposite irregular shape on the face of the head, the volume in the chamber can reduce to near zero, or whatever minimum volume is desired, just like with a traditional cylinder with flat piston and head surfaces, such that the same compression ratios can be accomplished with the same stroke lengths.

As used herein, the terms “head” and “housing head” and “cylinder head” are used to mean the stationary end portion of the compression chamber (e.g., the top end in the orientation of FIG. 1A) that is opposite from the piston head, and which may or may not be an integral portion of the overall stationary portion of the compressor. The head and the lateral side walls of the stationary portion of the compressor can, along with other stationary components, be referred to collectively as the “housing” of the compressor. The stationary lateral side surfaces of the compression chamber may or may not have a circular cross-sectional shape. For example, any cross-sectional shape (e.g., circular, elliptical, polygonal, etc.) can be provided for the lateral side surfaces of the compression chamber so long as that shape is consistent along the axial path of the piston head so that a seal can be maintained between the perimeter of the piston head and the lateral side surfaces of the compression chamber.

At the top of the piston stroke (sometimes referred to as “top dead center”), it can be desirable for the piston and head to approach very closely, with little space between them, so that the gas that has been compressed can be effectively cleared from the compression chamber. It is conventionally believed that surface features on the working side of the piston head and/or on the opposing housing head will reduce the potential compression ratio and clearance volume because they inhibit the piston and head from coming into very close proximity at the top of the stroke. However, as described herein, a near zero (or desirably small) compression chamber volume can still be achieved even when the working side of the piston head and/or on the opposing housing head include large projections extending into the compression chamber.

Thus, a piston-type compressor can be designed with many different shapes of piston and head, even shapes that are far different from the nearly flat to completely flat shapes that are conventional. And, the chamber side walls can be shapes other than purely cylindrical. This freedom allows inclusion of increased cooling surfaces that are disclosed herein. Piston and head geometries are disclosed that extend cooling surfaces into the inner, central regions of the compression chamber, and yet allow for close approach of the piston and head at top dead center, which makes it possible to provide piston-type compressors that compress a gas, while at the same time extract significantly more heat from the gas than in a conventional design, and thereby keep the gas from rising significantly in temperature due to compression heating.

In some of the embodiments disclosed herein, the lateral walls of the housing that form lateral surfaces of the compression chamber serve as enhanced heat extraction devices, shaped and placed to maximize the extraction heat from the compressing and exiting gas. The round, smooth side walls of a traditional cylindrical compressor extract a minimal amount of heat from the compressing and exiting gas due to their minimal surface area to gas volume ratio. The minimal loss of heat from the gas to the smooth side walls of a conventional cylindrical chamber is incidental to the large rise in temperature in the gas. By contrast, in disclosed compressor embodiments herein, the compression chamber walls are much more effective at extracting heat from the gas and thereby interfering with the temperature rise in the gas. For example, in some embodiments, projections extending axially from the piston head and the opposing housing head can also be integral with and/or in direct contact with the lateral side walls of the chamber, allowing them to efficiently conduct heat from all parts of the chamber, including the central parts of the chamber, both in the axial directions to the piston and head, but also laterally (e.g. radially) to the side walls of the housing. In some embodiments, for example, the projections can have major dimensions in the axial and radial directions (the desired directions of heat conduction), while being relatively thin in the circumferential or angular directions to provide a maximum surface area per volume ratio for the projections. The opposing piston and head projections can also be tapered in the axial direction so that they interleave and nest with each other to provide a minimal chamber volume at the top of the piston stroke.

The concept of incorporating increased cooling surface area in the compression chamber via projections that extend into the inner regions of the compression chamber and extract heat from throughout the compressing gas, without interfering with the compression, applies equally well to all types of compressors that involve compression chambers, such as screw compressors, wobble plate-type compressors, and scroll-type compressors. But, for the present discussion, the application of these concepts in a reciprocating piston-type compressors will be used as a representative, non-limiting example for ease of description.

Broadly described, exemplary devices for compressing gases and other compressible fluids can include a compression chamber of varying volume so that a compressible fluid can be drawn into the chamber as the volume of the chamber increases and can be compressed to a higher pressure and expelled from the chamber at a higher pressure as the volume of the chamber decreases. They can further include cooling surfaces that extend from the walls of the compression chamber, whether stationary or not, that define the compression chamber, into the inner regions of the compression chamber and extract heat from the compressing fluid in the inner regions of the compression chamber during the compression, without significantly hindering the variability of the compression chamber volume or substantially hindering the compression and expulsion of the fluid. The cooling surfaces can also be coupled to effective external cooling such that the heat extracted from the compressing and exiting fluid is subsequently efficiently extracted from the compression chamber, keeping the cooling surfaces from rising significantly in temperature, so that a large portion of the heat is extracted from the compressing fluid during each compression cycle.

Many different geometries consistent with the herein described improvements substantially improve heat extraction from the gas, and all such geometries are included and encompassed by this disclosure. Such geometries may include single cooling surfaces, multiple cooling surfaces that are arranged in rows, circles, other closed or annular shapes, spirals, are arranged radially or in arrays, are arranged symmetrically within the compression chamber, or are without significant symmetry or even significant pattern. Various different geometries can offer a combination of different advantages. Example advantages can include the following:

(a) Some geometries offer greatly increased surface area, such as additional surface expanses that the compressing and/or exiting gas can come into contact with and lose heat to.

(b) Some geometries are more effective than others in establishing effective flow paths for heat that is extracted from the compressing or exiting gas and conveyed to the external cooling sources.

(c) Some geometries, both symmetric and asymmetric, move some of the gas from one or more regions of the compression chamber to one or more other regions of the compression chamber during the compression and/or expulsion of the gas, resulting in greater flow of the gas within the compression chamber during compression and/or expulsion of the gas.

(d) Some geometries create an advantageous pattern or patterns to the flow of the gas as it enters the compression chamber, or during the compression and/or expulsion of the gas, patterns such as swirling or high-turbulence patterns of flow, that result in increased thermal contact between the gas and the walls of the chamber and/or the cooling surfaces.

(e) Some geometries create more turbulence in the flow of the gas over the compression chamber walls and/or the cooling surfaces during the compression and/or expulsion of the gas, and thus improve the exchange of heat between the gas and the walls of the compression chamber and/or the cooling surfaces; as examples, some or all of the surfaces of the compression chamber walls and cooling surfaces can be undulating, furrowed, dimpled, rough, or polished.

There are other advantages, besides these few examples, that can be achieved through various geometries of the walls of the compression chamber and the cooling surfaces. All such geometries and advantages are included in this disclosure and encompassed by the claims of this application where applicable.

Continuing with the example of the application of this technology to a piston-type compressor, FIG. 1 shows one of many exemplary cooling surface geometries that can be incorporated in a compression chamber. In FIG. 1, the upper geometry, including arrayed projections 24, can be stationary and located in the stationary upper end of the chamber and optionally integral with the head and/or side walls of the housing. The lower geometry, including arrayed projections 34, can be movable and optionally integral with the piston head. For purposes of illustration, the opposing cooling surfaces are shown separated more than their maximum spacing when the piston is at the bottom of the stroke. This exemplary geometry is circularly symmetric (a.k.a., rotationally symmetric) with twelve projections on each opposing portion and a twelve-fold (C12) symmetry. In other embodiments, the geometry can include 4, 6, 8, 10 or other numbers of projections (including odd numbers) on each opposing section, and corresponding different levels of symmetry. Also, it is not necessary that a piston, chamber, and head have circular symmetry, or any symmetry at all.

In any of the compressor embodiments disclosed herein, the described geometry can significantly increase the surface area of the compression chamber compared to a conventional cylindrical chamber having flat top and bottom walls. In comparison to a conventional cylindrical chamber having flat top and bottom walls (taking for comparison a cylindrical chamber wherein the vertical height of the cylinder is equal to the diameter of the cylinder), a compression chamber having geometry as described herein and having an equivalent diameter and height (vertical distance between opposing top and bottom surfaces) can have a surface area that is at least 200%, at least 300%, at least 400%, and/or at least 500% greater. The particular example geometry shown in FIG. 1 has more than five and a half times the surface area of an equivalent cylinder with a flat piston and a flat head. This increase in surface area, optionally along with any combination of other cooling features disclosed herein, results in an increase in the head extraction from the compressing gas and a corresponding increase in efficiency of the compressor. For example, for the disclosed compressors, the energy efficiency of the compressor can be greater than 33%, greater than 50%, greater than 66%, and/or greater than 75%.

In this example geometry, the cooling surfaces also have two optional features: (1) the opposing surfaces of the housing head and of the piston head are identical; and (2) they also “nest” perfectly. That is, if brought together all the way to the point of contact, there is no space between the opposing surfaces. The ability to nest perfectly makes higher compression ratios (lower clearance volumes) possible; and, efficient chamber designs may tend in this direction. But, in some cases, perfect nesting may well be sacrificed to some extent, in order to capture other advantages of geometries that do not evince perfect nesting.

Another advantage of the exemplary geometry of FIG. 1, and many others, is that the cooling surfaces of the head can be continuous with (i.e., part of or integral with), or contiguous with and in intimate thermal contact with, the side walls of the housing, which can itself be cooled externally (e.g., with cooling fins, a water jacket, fluid spray, and/or other cooling mechanisms). Whether the projections in the compression chamber are continuous/integral with or contiguous to the housing walls or not, the projections can convey heat to and lose heat to external cooling that is applied alongside of the lateral housing walls that form the compression chamber. The heat extraction, therefore, is not limited to extraction through the head and piston only. This greatly increases the efficiency of removing heat from the gas as it compresses and as it is expelled from the compression chamber.

A useful effect of this exemplary geometry, and many other possible geometries, is exchange of heat between the stationary and the non-stationary walls of the compression chamber. Due to differences in the ability to cool the stationary and non-stationary walls of the compression chamber, it can be more difficult to keep the temperature of the non-stationary wall(s) of the compression chamber (e.g., the piston head in a piston-type compressor) as low as desired. With the exemplary geometry of FIG. 1, and many others, the (possibly turbulent) flow of gas into the compression chamber during the inlet phase of the cycle can extract heat from higher temperature wall portions of the compression chamber, and can release heat to cooler temperature wall portions of the chamber as the gas enters the chamber. Heat is thus transferred between the different walls of the compression chamber, tending to even out the temperatures of the compression chamber walls. This effect can prove especially useful in keeping the non-stationary wall(s) of the compression chamber cooler, as they can, in some geometries, prove to be the more difficult walls to cool. To enhance this effect, it is beneficial for the compression chamber geometry to increase the flow of gas within the chamber (intra-chamber flow) and increase the turbulence of the gas as it moves in the compression chamber, which thereby increases heat transfer between the walls and the gas.

Cooling Moving Walls with Cooling Fluid

The stationary walls of a compression chamber are relatively easy to provide with sufficient external cooling to keep the stationary surfaces of the chamber at low enough temperatures to continuously extract heat from the gas over many compression cycles. However, a non-stationary wall of a compression chamber, such as the piston head in the present example of a piston-type reciprocating compressor, can be more difficult to provide with effective cooling. In piston-type compressors described herein, the piston head is used not only as a compressing surface, but also as an active member in extracting heat from the gas. The heat that the piston head extracts (or accepts) from the gas must be carried away and delivered to an effective heat sink, in order to keep the piston from eventually heating up and becoming ineffective as a cooling surface.

Generally, a non-stationary wall of a compression chamber cannot be cooled with a stationary water jacket or with stationary external cooling fins, as in the case of a stationary walls, because there is no pathway to efficiently convey heat from the moving compression chamber wall to a heat sink such as a water jacket or a stationary cooling fin. However, in some embodiments disclosed herein, a moving heat conveyance pathway can be provided to carry the heat extracted from the compressing gas by the non-stationary compression chamber wall to one or more, possibly stationary, external heat sinks. In some embodiments, an intermediary substance or object is provided that can both extract (or accept) heat from a non-stationary compression chamber wall and convey it to and release it to a, possibly stationary, heat sink. Some such embodiments utilize a fluid that can serve as the intermediary heat transfer substance. In some embodiments of a piston-type compressor, oil cooling of the non-working side of the piston can convey the heat extracted by the piston from the compressing and exiting gas to an external heat sink, such as, for example, a radiator. Cooling fluids other than oil can alternatively be used. In some embodiments, the cooling fluid can also act as a lubricant, such as using the same fluid that is used to cool the non-stationary compression chamber walls. As used herein, the term “fluid” includes any combination of gas, liquid, vapor, suspension, or other substance that flows under applied shear stress.

The use of a cooling fluid (e.g., oil) conducted into contact with the non-working side of the piston head, or other moving compression chamber wall, serves the purpose of extracting heat from the compressing gas, through the piston head, to a remote heat sink. This is fundamentally different than the use of oil to cool a piston in an internal combustion engine, where the purpose of the cooling is to keep the piston itself from getting extremely hot and melting or otherwise failing as extreme heat is generated in the chamber from fuel explosions. In that context, any rudimentary cooling means can be sufficient due to the extreme temperature gradient between the exploding fuel and the ambient surroundings. By contrast, in the disclosed compressors, it is desired to keep the low temperature gas in the chamber from heating up in the first place, and therefore the cooling fluid is used to quickly remove any heat that the piston head extracts from the compressing gas so the piston does not heat up and remains at or near ambient temperature or otherwise low temperature. In fact, cooling of a piston in an internal combustion engine to near ambient temperature is very undesirable, because the piston would extract too much heat from the chamber (high chamber temperature is good in the engine context) and would reduce the efficiency of the engine and constitute over-cooling of the piston.

Broadly described, exemplary devices for compressing gases and other compressible fluids can include a non-stationary wall of a compression chamber that in moving varies the volume of the compression chamber, causes or corresponds to gas being drawn into the chamber, causes the gas to be compressed to a smaller volume and higher pressure, and causes the gas to be expelled from the chamber at or near the higher pressure, and the non-stationary wall of the chamber is effectively cooled by occasional, continual, or continuous contact with a cooling fluid of lower temperature to which it can lose heat and thereby remain as close as possible to the ambient temperature, or to be cooled to a temperature below ambient.

Fluid cooling of non-stationary compressor chamber walls applies to all types of compressors, not just piston-type compressors. Further examples include: (1) The non-stationary scroll in a scroll-type compressor can be effectively cooled through contact with oil or another fluid on its non-working side; (2) The counter-rotating rotors of screw-type compressors and roots-type blowers can be effectively cooled by pumping water or another fluid through them; and (3) The diaphragm of a diaphragm-type compressor can be cooled by contact with a cooled fluid on its non-working side.

Increased Surface Area on Non-Working Side of Moving Wall

It can be difficult to extract sufficient heat from a piston or other moving chamber wall due to the relatively low surface area on the non-working side of the piston head or other moving chamber wall. For example, the non-working side of a conventional air compressor piston is lacking in any cooling fins or other irregular features that would increase surface area and therefore heat loss to oil or other cooling fluid that they might be in contact with, either occasionally, continually or continuously.

In order to serve a significant role in extracting heat from the gas during compression, as described herein, the non-stationary wall(s) of the compression chamber can be maintained at a low enough temperature (e.g., near ambient) to be effective in extracting heat from the compressing and exiting gas. In order for the non-stationary wall(s) to be maintained at a low enough temperature, in some embodiments, the non-working side of the non-stationary wall(s) can be structured to provide increased surface area and to effectively transfer heat to the cooling fluid that contacts it. For example, the non-working side of a piston head can include fins, ribs, ridges, grooves, projections, bumps, recesses, and/or other non-flat surface shapes that increase the surface area, such as to a surface area substantially greater than the cross-sectional area of the bore in which the piston head reciprocates. Such a surface structure can increase the area that contacts the cooling fluid to increase the amount and rate of heat transfer from the piston head to the cooling fluid. Examples of such piston head geometry are shown in the exemplary embodiments included herein.

Broadly described, exemplary devices for compressing gases and other compressible fluids can include a non-stationary (moving) wall of a compression chamber that in moving varies the volume of the compression chamber and causes a gas to be drawn into the chamber and causes the gas to be compressed to a smaller volume and a higher pressure and to be expelled from the chamber at a higher pressure, the non-stationary wall of the chamber being designed and structured to incorporate one or more surfaces (on the non-working side of the non-stationary wall) that are not in contact with the compressible fluid being compressed, said surface(s) increasing or maximizing the area of the non-working side of the non-stationary (moving) wall, the effect of which is to increase the ability of the non-stationary wall to lose heat from its non-working side to a cooling fluid, said surface(s) being in occasional, continual, or continuous contact with a cooling fluid of lower temperature to which they can lose heat.

Non-Axial Motion During Axial Compression

With the example geometry in FIG. 1, the gas flows over the cooling surfaces to some extent during compression, and much more during expulsion of the gas from the compression chamber. This is true of many of the other possible geometries of the chamber, head, and piston, to a greater or lesser degree. However, with the example geometry of FIG. 1 and many of the other geometries of the chamber, head, and piston, during the majority of the compression of the gas, flow of the gas over the surfaces of the compression chamber may be limited. Given the insulating properties of stagnant or slow moving gas, the gas may rise more in temperature than would be desirable, at least in some areas of the compression chamber. In this sense, the compression can be partially, locationally, adiabatic, and this increases the work required to compress the gas. In the example geometry of FIG. 1, and many of the other possible geometries, the excess heat generated from this partially adiabatic compression subsequently is at least partly lost to the cooling surfaces as the gas exits the compression chamber, because during that part of the cycle (expulsion) the gas develops relatively higher velocities and greater turbulence. And, this later loss of excess heat to the cooling surfaces is helpful in reducing the total amount of work required to compress the gas and force it out of the compression chamber. As the gas loses heat to the cooling surfaces, the pressure of the gas relaxes considerably, easing the final stages of the compression. Still, even though the excess heat generated by the partially adiabatic compression is subsequently partially removed before and during the expulsion of the gas from the compression chamber, and even though the gas exits the compression chamber at a temperature much closer to the temperature of the gas as it entered the compression chamber, the temperature did rise during compression, and excess work was therefore done, which decreases the overall efficiency of the compressor. To minimize adiabatic compression of the gas, greater flow, and especially more turbulent flow, of the gas over the cooling surfaces during the compression is advantageous and desirable.

Augmented flow, and especially turbulent flow, of the gas during compression can be achieved, in some embodiments, by shifting a portion of the gas from one or more areas of the compression chamber to one or more other areas (or between areas) of the compression chamber during the compression. That is, flow in and about the compression chamber that does not involve expulsion of the gas can increase the transference of heat from the gas to the cooling surfaces during the compression. In the example of a piston-type compressor, this can be achieved by manipulating the head, side walls, and/or the piston to provide a stroke that is other than the direct, straight-line axial stroke typical of many piston-type compressors. For the example geometry of FIG. 1, and many of the other possible geometries, one way to shift the gas in and about the compression chamber (intra-chamber flow) is to rotate, shift, or oscillate the head and/or the piston slightly, in a non-axial direction, during the axial compression motion of the piston. Such non-axial motion can, for example, be oscillatory (back and forth) or in a single direction. There are many other non-axial motions or manipulations of the head and/or piston that can be advantageous for increasing flow and turbulence in the chamber, including in other exemplary compressors, such as wobble-piston compressors and swash plate-driven compressors, both of which can involve a tipping-back-and-forth motion of the piston (among other non-axial motions) as it reciprocates axially in the cylinder(s) of such compressors.

In some embodiments, the housing head and/or side walls can be oscillated or otherwise moved during the straight line axial stroke of the piston, instead of oscillating the piston itself. Or, both the piston and the housing can be moved, such as in opposite directions, to create a relative oscillatory motion between the piston and the housing.

FIG. 12 shows a rotationally oscillating path of approach for the example geometry similar to that shown in FIG. 1, in which one or more of the housing head, side walls, and the piston are manipulated to create a rotational oscillation about a central axis. FIGS. 11A-11C shows the non-axial rotation motion in isolation from the axial, straight-line reciprocating approach of the head and piston that is typical of many piston-type compressors. The side-to-side relative motion reduces some volumes and increases or creates other volumes between the interleaving projections, causing the gas to flow along the surfaces. The flow is largely across or parallel to the faces of the cooling surfaces, which enhances heat transfer into the projections. Turbulence during such flows can be increased by detailed sculpting of the surfaces. Undulating surfaces, for example, can be included in some embodiments to increase turbulence. Such detailed sculpting is omitted from FIGS. 11 and 12 for clarity. As a cooling surface of the piston (or head) makes a near approach to its adjacent cooling surface on the head (or piston), higher velocity flow (sometimes referred to as “squish”) is developed, creating significant and useful additional turbulence in the gas, some at points distant from the near-contacting surfaces. Flow rates of up to a few hundred miles per hour can be achieved, for example, and turbulence of any practically significant level is readily attainable. And, as the opposing surfaces then move apart again, a considerable low-pressure point is created between the separating surfaces, causing what might be termed “back-squish”. This also can be highly turbulent flow, further enhancing heat transfer to or from the gas.

In addition to the exemplary oscillatory path shown in FIG. 12, various alternative paths and orientations can be achieved in other embodiments, and many would be as effective or more effective in actual practice.

During both the inlet phase and during the compression/exhaust phase, the increased flow and turbulence caused by the non-axial motion of piston and/or head can help maintain more even temperatures in the gas and in the walls of the compression chamber.

Broadly described, exemplary devices for compressing gases and other compressible fluids can include one or more compression chamber walls that move (e.g., oscillate) in any motion that is different from the normal motion that is characteristic of that type of compressor (such as moving non-axially during an axial compression stroke of reciprocating piston-type compressor, or non-orbital motion during the normal orbiting motion for a scroll-type compressor, and equivalent non-normal motions for a blower type compressor, screw-type compressor, wobble-piston compressor, or a swash plate-driven compressor), an effect of which non-normal movement is to create or increase intra-chamber flow of the compressing gas and/or increase turbulence in the gas.

Exemplary Embodiments

FIGS. 1A and 1B show an exemplary reciprocating piston-type compressor 10 having many of the features disclosed herein. The compressor 10 generally includes a stationary upper housing 12, a reciprocating piston 20 that forms a sealed compression chamber along with the housing, and a lower casing 13 that can comprise a crank case for driving the piston and a cooling fluid reservoir. This disclosure encompasses compressor embodiments having any number of such compression chambers.

The housing 12 includes an upper housing head 14, side walls 18, and other stationary components. A piston 20 reciprocates within the housing side walls 18 using annular seals 22 (e.g., O-rings) to create a sealed internal compression chamber 16. The side walls 18 and piston 20 can have any cross-sectional bore profile, such as cylindrical or otherwise. The material of the housing 12 and the piston 20 can comprise any sufficiently strong, thermally conductive material (e.g., any of various metals) that facilitate heat conduction away from the gas in the compression chamber. The stationary components, including the head 14 and side walls 18, can be one-piece/integral with each other or can be comprised of components that that are contiguous and in sufficiently close contact with each other to transmit heat across the interfaces without substantial thermal insulation therebetween. The stationary components can, for example, be glued, soldered, welder, brazed, or bonded together.

The stationary head 14 includes projections or protuberances 24 that extend into the compression chamber 16 and the piston 20 includes corresponding opposing projections or protuberances 34 that extend into the compression chamber. The projections 24 and 34 partially divide the compression chamber 16 into smaller spaces, such that a substantial fraction of the gas in the chamber is at all times relatively close, compared to an open compression chamber, to one or more thermally conductive surfaces to which the compressing gas can give up heat. The projections 24 and 34 can have various geometries, with one example shown in FIG. 1. A further example of such geometry is shown in FIGS. 4-9.

As illustrated in FIG. 1, the opposing projections 24 and 34 can be shaped to interleave or mesh with each other to avoid interfering with each other at the top of the compression stroke, as the piston pushes the lower projections 34 up into and between the upper projections 24. At the extremity of the compression stroke (top dead center), the lower projections 34 can approach the upper projections 34 so as to define a minimum volume in the compression chamber that is a small fraction of the working volume (swept volume) of the compressor, thus resulting in the compressor having a relatively small clearance volume and high compression ratio. The lower projections 34 may or may not actually contact the upper projections 24 at the top of the stroke.

The projections 24 can be integral with the rest of the head 14 and the lateral/radial surfaces of the projections 24 can be integral with or in thermally conductive contiguous contact with the side walls 18, since the projections 24, the rest of the head 14, and the side walls 18 can all be (not necessarily) stationary. This allows heat from the projections 24 to be conducted both axially up through the head 14 (as illustrated by arrows 26) and laterally/radially through the side walls 18 (as illustrated by arrows 28). The lower axial ends of the projections 34 can be integral with the rest of the piston 20 while the lateral/radial surfaces of the projections 34 can be in sliding contact or near-contact with the stationary side walls 18, such as with a lubricant therebetween. This allows heat from the projections 34 to be conducted both axially down through the piston 20 (as illustrated by arrows 36) and laterally/radially through the side walls 18 and sealing rings 22.

As shown, the projections 24 and 34 can be gradually tapered and can have a truncated flat distal end, such as with slight rounded edges, or can have smoothly rounded distal end. The projections can be, though not necessarily, arrayed angularly around a central vertical axis, with each projection extending radially away from the central axis in a generally wedge or trapezoid shape. The radially outward facing surfaces can be rounded to match the side walls 18 of the housing, thereby facilitating heat flow in the radial direction through the projections and into the surrounding side walls. The projections can have broad, flat angularly-facing sides that form valleys between them, the valleys having the same or similar size and shape as the projections, but axially flipped, such that the valleys can fittingly receive the opposing interleaving projections at the top of the compression stroke. The projections 24 of the housing head 14 can angularly surround an open central bore that couples the intake 50 and exhaust 52 with the compression chamber 16. Each of the open valleys between the projections 24 can extend radially from the central bore. The lower projections 34 of the piston head can also include/surround a corresponding central open area. The geometry of the projections 24 and 34 ensures that the gas in the compression chamber 16 is always close to at least one surface that is ready to accept heat from the gas as it is compressed.

The compressor 10 includes a gas intake 50 and exhaust 52 that are fluid coupled to the compressor chamber via at least one valve 48. The intake and exhaust can flow through a common upper central passage where the valve 48 is located. The valve can be mounted in the housing head 14 and/or in another component of the housing 12.

The stationary housing 12 can include a cooling jacket 30, external fins, and/or other features that accept heat from the head 14, side walls 18, projections, and other components and conduct the heat away from the compressor. In the illustrated embodiment, the cooling jacket 30 (which can utilize any cooling fluid) extends around a substantial portion of the side walls 18 and above the head 14. The cooling jacket 30 includes a lower outlet 62 coupled to a heat sink 32 via conduit 54, and an upper inlet 64 coupled to the heat sink via conduit 56. The cooling fluid cycles through the cooling jacket 30 and the heat sink 32 to continuously remove heat from the compressor. The heat sink 32 can comprise any configuration sufficient to cool the cooling fluid below the temperature of the housing 12 so that the fluid can effectively extract heat from the housing. Fins, fans, and/or other means can also be included with the housing 12 to help remove heat from the outer surfaces of the housing walls.

The motion of the reciprocating piston 20 can be driven by a crank shaft located in the casing 13 and coupled to the piston head with a piston shaft 60. The piston shaft 60 is pivotably joined to the head of the piston 20 such that the head of the piston moves up and down axially while the shaft 60 moves in a more complex motion.

The piston 20 has a working side that includes the projections 34 and which forms surfaces of the compression chamber, and a non-working side facing away from the compression chamber. Heat generally flows from the working side through the piston head to the non-working side, and the heat is then conducted away from the non-working side to a heat sink. The non-working side of the head of the piston 20 includes features that increase its surface area, such as fins 40, to enhance heat conduction away from the piston to a cooling fluid, such as air and/or a cooling fluid. The non-working side can also or alternatively include other surface features, such as ribs, grooves, projections, bumps, etc., that increase the surface area and/or facilitate the flow of a cooling fluid over the surface to effect heat extraction from the piston.

The compressor 10 can include a piston cooling system that utilizes a cooling fluid that is cyclically conducted into contact with the non-working side of the piston 20 and then conducted to a heat sink. The piston cooling system can include a cooling fluid reservoir 42, such as located within the casing 13, a pump 58 to conduct the cooling fluid through a conduit and cause the cooling fluid to be sprayed from an nozzle 44 in the direction of the non-working side of the piston 20 and/or the piston shaft 60, and an external heat sink 46 coupled to the reservoir 42 and pump 58. Alternatively, the cooling fluid can be circulated through a portion of the head and/or lateral walls of the compressor and thereby lose heat to the cooling mechanism(s) provided for the head and/or lateral walls. This can eliminate the need for a separate heat sink for the piston cooling fluid. The dashed lines in FIGS. 1A, 1B, 2A, and 2B emanating from the nozzles/sprayers 44 and 144 illustrate exemplary spray paths for the cooling fluid, however many other pray patterns can be achieved, causing the fluid to contact any combination of the piston surfaces. The cooling fluid can be sprayed in such a way to effect substantial contact of the fluid with the large surface area of the non-working side of the piston. For example, if fins 40 are present, the cooling fluid can be sprayed onto the fins and into the valleys between the fins to cover as much of the surface area of the fins as possible. The cooling fluid can be allowed to run via gravity along the surfaces of the piston and/or fall down toward the reservoir, taking with it heat from the piston. The relatively hotter cooling fluid is then pumped through an external heat sink 46, or through a portion of the head and/or walls, to be cooled down and then pumped back through the cycle to be sprayed at the piston again to help maintain the piston at or near ambient temperature. In some embodiments, more than one spray nozzle or other spraying device can be used to maximize the volume and coverage of the fluid in contact with the piston at any given time.

In some embodiments, one or more of the heat sinks used to remove heat from cooling fluids (e.g., heat sink 32 and/or heat sink 46) can comprise a well or other subterranean cavity containing water that is below the ambient air temperature, or can comprise a fluid cooled to below ambient air temperature of the air surrounding the compressor by colder outdoor air during colder times of the year.

FIGS. 2A and 2B show another exemplary reciprocating piston-type compressor 100 that is similar to the compressor 10, but also includes means for causing the reciprocating piston to oscillate non-axially (e.g., rotationally about its central axis) while reciprocating along the central axis as usual. The compressor 100 generally includes a stationary upper housing 112, a reciprocating piston 120 that forms a sealed compression chamber along with the housing, and a lower casing 113 that can comprise a crank case for driving the piston and optionally a cooling fluid reservoir.

The housing 112 includes an upper housing head 114, side walls 118, and other stationary components. A piston 120 reciprocates within the housing side walls 118 using annular seals 122 (e.g., O-rings) to create a sealed internal compression chamber 116. The side walls 118 and piston 120 have a cylindrical cross-sectional bore profile to allow for rotational oscillations of the piston, though in alternative embodiments the oscillatory motion of the piston can be other that rotational and thus the cross-sectional bore profile can be other than cylindrical.

The stationary head 114 includes projections or protuberances 124 that extend axially into the compression chamber 116 and the piston 120 includes corresponding projections or protuberances 134 that extend axially into the compression chamber. The projections 124 and 134 partially divide the compression chamber 116 into smaller spaces, such that a substantial fraction of the gas in the chamber is at all times relatively close, compared to an open compression chamber, to one or more thermally conductive surfaces to which the compressing gas can give up heat. The projections 124 and 134 can have various geometries, with one example shown in FIG. 1 described elsewhere herein.

The compressor 110 includes a gas intake 150 and exhaust 152 that are fluid coupled to the compressor chamber via at least one valve 148. The intake and exhaust can flow through a common upper central passage where the valve 148 is located. The valve can be mounted in the housing head 114 and/or in another component of the housing 112.

The stationary housing 112 can include a cooling jacket 130, external fins, and/or other features that accept heat from the head 114, side walls 118, projections, and other components and conduct the heat away from the compressor. The cooling jacket 130 includes a lower outlet 162 coupled to a heat sink 132 via conduit 154, and an upper inlet 164 coupled to the heat sink via conduit 156. The cooling fluid cycles through the cooling jacket 130 and the heat sink 132 to continuously remove heat from the compressor. The heat sink 132 can comprise any configuration sufficient to cool the cooling fluid below the temperature of the housing 112 so that the fluid can effectively extract heat from the housing. Fins, fans, and/or other means can also be included with the housing 112 to help remove heat from the outer surfaces of the housing walls.

The motion of the reciprocating piston 120 can be driven by a crank shaft located in the lower casing 113 and coupled to the piston head with a piston shaft 160. The motion of the piston head is more complex than simple axial reciprocation along the central axis of the cylindrical housing bore. For example, the piston head can move rotationally, or angularly, within the bore during the axial compression stroke. In some embodiments, the piston can tilt or wobble, such as about an axis perpendicular to and/or intersecting the reciprocation axis) during reciprocation. The effect of such complex motion is to increase flow and turbulence of the gas in the compression chamber, especially during compression when heat is being generated, as is described in more detail elsewhere herein.

In some embodiments, the piston can complete at least one full oscillation during each compression stroke. A full oscillation includes motion in one rotational direction followed by motion in the opposite rotational direction such that the piston returns to its original rotational position. The piston may complete more than one oscillation during the stroke in some embodiments, such as at least two oscillations. In some embodiments, the piston may complete a half oscillation during the compression stroke, and complete the rest of the oscillation during the down stroke, such that one full oscillation is completed in one full reciprocation.

An exemplary axial and rotational path 170 of the distal tip of one of the lower projections 134 is shown in dashed lines in FIGS. 2A and 2B. In this example, one full oscillation is completed per reciprocation cycle of the compressor. This exemplary path includes a portion between point 172 at the bottom of the stroke and point 174 part way through the compression stroke, wherein the projections 124 and 134 approach each other and create intra-chamber flow of the gas from a shrinking space on one side of the projections to a space on the other side of the projections, thus creating increased flow velocity and turbulence. As the opposing surfaces approach at point 174 without touching, “squish” occurs and causes rapid flow of the gas being squeezed out from between the approaching surfaces, creating higher velocity and increased turbulence. The path can then include a portion between points 174 and 176 where the opposing projection surfaces, having approached one another closely at 174, now separate to a greater distance, resulting in “back-squish,” wherein the gas rapidly flows back into the low pressure void growing between the surfaces. This again generates increased gas flow and turbulence. The path can further include a portion between points 176 and 178 where the projections 134 are aligned in the valleys evenly between the projections 124 such that the piston moves in a substantially straight line motion (e.g., only axially) to allow tight intermeshing of the opposing projections without contact between them. This motion is reversed during the down stroke.

An alternative path 300 is illustrated in FIG. 12, wherein a piston completes more than one complete oscillation during the compression stroke. The path 300 shown in FIG. 12 is not necessarily due to rotational motion of the piston, as other types of non-axial motion (e.g., in non-cylindrical bore shapes) can result in the path shown in FIG. 12 as well. As shown in FIG. 12, the lower projection 302 moves right, then left, then right and finally back to a central aligned position at top dead center. This motion may or may not be reversed during the down stroke.

FIGS. 11A-11C illustrate the induced flow of the gas caused by the non-axial relative motion between the opposing projections of the piston and the head. The illustrations of FIG. 11A-11C ignore the actual axial motion occurring simultaneously between the opposing projections in order to more clearly describe the effects of the non-axial motion. In FIG. 11A, the upper projections 300 are shifted to the right relative to the lower projections 302, causing opposing surfaces 304 and 306 to come into close proximity and causing gaps 312 between opposing surfaces 308 and 310. In FIG. 11B, the top projections have shifted partially to the left relative to the lower projections 302, causing the gaps 312 to be reduced and opening gaps 320 between the surfaces 304 and 306. This causes the gas to flow from the gaps 312, around the valleys 314 and 316, into the gaps 320. In FIG. 11C, the gaps 312 are substantially closed and the gaps 320 are increased as the gas continues to flow around the valleys 314, 316 and into the gaps 320. This flow of gas can be at high speed due to the “squish” effect and therefore create desirable and substantial turbulence. The flow is also primarily directed parallel to and close to the surfaces of the projections, which maximizes heat transfer between the walls and the gas.

Referring again to FIG. 2A, an exemplary mechanical means for causing the illustrated path 170 is shown. The inner walls of the housing 12 can include one or more grooves 180, and the outer wall of the piston 20 can include one or more corresponding pins 182 that extend radially into the grooves and move along the grooves as the piston moves axially. Or, this can be reversed, with the piston having the pins and the cylinder walls having the grooves. The shape of the grooves 180 defines the non-axial motion of the piston 20 and the corresponding path of the lower projections 134. In the example of the compressor 100, each groove 180 has a shape that matches the shape of the path 170.

The piston shaft or connecting rod 160 can include a thrust bearing 184 that allows the piston head freedom to move in oscillatory motions in addition to its normal reciprocation motion while the lower portion of the piston shaft 160 moves in its normal elliptical/reciprocating motion.

The non-working side of the head of the piston 120 can include features that increase its surface area, such as fins 140, to enhance heat conduction away from the piston to a cooling fluid, such as air and/or a cooling liquid. The non-working side can also or alternatively include other surface features, such as ribs, grooves, projections, bumps, etc., that increase the surface area and/or facilitate the flow of a cooling fluid over the surface to effect heat extraction from the piston.

Similar to the compressor 10, the compressor 100 can also include a piston cooling system that utilizes a cooling fluid that is cyclically conducted into contact with the non-working side of the piston 120 and then conducted to a heat sink. The piston cooling system can include a cooling fluid reservoir 142, such as located within the lower casing 113, a pump 158, one or more nozzles 144 to spray the fluid in the direction of the non-working side of the piston 120 and/or the piston shaft 160, and an external heat sink 146. In some embodiments, more than one spray nozzle or other spraying device can be used.

FIGS. 3-10 show various views of an exemplary compressor 200 that includes many of the features disclosed herein. FIG. 3 shows an assembled view and FIG. 4 shows an exploded view. The compressor 200 includes a base 202, main body 204, outer sheath 206, inner liner 208, piston 210, head 212, valve housing 214, and/or various other components. All of the illustrated components form a stationary housing, except for the piston 210, which reciprocates within the housing and can be driven by any suitable power source (not shown). The piston 210 and head 212 include opposing, interleaving projections similar to those described with compressors 10 and 100.

The compressor 200 can include an intake/exhaust port 216 at the center top of the head 212 and valve housing 214, which can include a valve to selectively allow gas intake or gas exhaust.

The compressor 200 can also include a drive mechanism (not shown) for causing the piston reciprocation, and can include a lower casing that houses the drive mechanism and/or includes a fluid cooling system (e.g., a fluid reservoir and spraying mechanism for cooling the non-working side of the piston, as described elsewhere herein).

The housing can include a fluid cooling jacket in a space formed in the housing between the outer surface of main body 204 and the outer sheath 206, as previously described in other embodiments, and the cooling jacket can extend over the top of the compression chamber in a space within the head 212 or in a space between the head and the valve housing 214. The cooling jacket can be fluidly coupled to an external heat sink through one or more inlets and outlets, such as via port 218 (FIG. 3) in the base 202. The main body 204 can be secured to the base 202 such that the stepped lower end 218 of the main body seals against the lower surface 220 of the base, while the outer sheath 206 secures against the top of the base, creating a sealed internal void for cooling fluid to flow from the cooling jacket to the port 218.

The entire structure shown in FIG. 3 can be mounted to a crank case or other stationary lower structure that contains a power means for driving the piston 210 and/or a fluid cooling means for cooling the non-working side of the piston 210, such as is described herein with reference to compressors 10 and 100.

The inner liner 208 is optional and can be mounted inside the main body 204 to provide a cylindrical surface across which the piston slides, such as via one or more O-rings in a sealed manner. The inner liner 208 can comprise a different material than the main body 204, and/or can be replaceable when worn.

FIG. 5 shows the piston 210 and the head 212 fully mated together, illustrating the corresponding but opposite geometry of the opposing projections 228 and 230. Each projection can fit very tightly in the valley between two of the opposing projections, such that the volume of the compression chamber formed therebetween is minimal at top dead center and the compression ratio and swept volume are increased. Flow and turbulence of the compressed gas is also increased with such geometry, especially at the later stages of compression.

FIGS. 6-8 show the head 212 in various views, and FIGS. 9-10 show the piston in various views. The head includes tapered projections 230 arrayed angularly around a central opening 236, with evenly sized valleys 231 formed between the projections. Similarly, the working side of the piston 210 includes projections 228 arrayed around a central opening 252 and sized to fit in the valleys 231 of the head, and valleys 229 sized to receive the projections 230 of the head. The projections of the piston and the head can optionally include notched regions (for example as shown at 254 in FIG. 9) at their radially inner ends near the intake/exhaust port 216, such as to provide increased space for intake and exhaust flow.

The head 212 can also include an upper body that includes two concentric rims 232 and 234. The inner rim 232 can be secured to and contiguous with the top end of the main body 204 while the outer rim 234 can be secured to the top end of the outer sheath 206, allowing the cooling jacket to extend into the upper body of the head 212 and or between the head and the valve housing 214.

As shown in FIGS. 5 and 9, the piston can include one or more recesses 226 in its outer surface to accommodate O-rings or other sealing members that provide for sealed sliding against the inner liner 208. The piston 210 can also include openings 250, reinforced wall structures 264, and/or other features that facilitate the jointed attachment of a piston rod or other mechanical linkage.

As shown in FIG. 10, the lower, non-working side of the piston 210 can include a recessed volume, generally referred to as 260, that includes one or more fins 262 that project downward from the otherwise flatter surface 266 on the non-working side of the piston. The illustrated embodiment includes five fins 262 having straight and curved shapes, and formed valleys between them, which substantially increase the surface area on the non-working side of the piston and enhance heat conduction through and away from the piston. As discussed elsewhere herein, a fluid cooling system can be used to spray or otherwise conduct cooling fluid into contact with the non-working side of the piston 200 to assist in removing heat from the piston. The increased surface area on the non-working side of the piston and/or such a fluid cooling system for the piston, can help keep the piston at or near ambient air temperature, so that the piston can actively conduct heat away from the compressed gas before the temperature builds up in the gas.

FIGS. 13-15 illustrate another exemplary compressor 400 that embodies many of the features disclosed herein. The compressor 400 is a wobble plate compressor that includes a plurality of reciprocating piston-based compression chambers, such as chambers 402 and 404 (only two are shown for clarity), which embody many of the principles that are described in connection with other reciprocating piston-type compressors disclosed herein, along with other features.

FIG. 13 shows a partially cross-sectional view of the wobble plate compressor 400, with certain features not shown in full cross-section. For example, the opposing projections 420 and 422 in the compression chambers are shown in a view that illustrates the interleaving structure and orientation of the opposed projections, similar to the illustrations in FIGS. 1A, 1B, 2A, and 2B, all of which can represent various geometries of projections, including the geometries shown in FIG. 1 and in FIGS. 4-9.

The disclosed technology can also be applied in swash-plate compressor embodiments in similar manner. In a wobble plate compressor, as shown in FIG. 13, a tilted drive plate 406 (as opposed to a crank shaft in a reciprocating compressor) is rotated by the shaft 408 of the compressor. In this manner, the drive plate processes about its center line 410, and in this manner forces the pistons 412, through the connecting rods 414, to reciprocate in the cylinders 416. Thus, a wobble-plate compressor (as well as a swash-plate compressor) is a species of the reciprocating compressor genus, and can include any combination of the features disclosed herein that are associated with a reciprocating compressor.

In a conventional wobble-plate compressor, the pistons and the heads are flat, and the compression chambers are simple cylinders, thus minimizing the surface areas of the compression chamber walls, and reducing and perhaps minimizing turbulence in the cylinders. For clarity, only two pistons are shown. However, the disclosed wobble-plate compressors can have more pistons, including odd numbers of pistons. Three, five, and seven piston wobble plate compressors are exemplary configurations that can include the disclosed technologies. The disclosed wobble-plate compressors can be used, for example, in automotive air conditioning systems, refrigerators, and air conditioning systems for buildings.

In FIG. 13, the compressor 400 includes contoured surfaces that project into the compression chamber, including projections 420 that protrude from the piston and opposing offset projections 422 that protrude from the stationary end of the chamber. As with other reciprocating compressors disclosed herein, many other geometries are possible, and many of those other geometries may be more or less efficient or otherwise advantageous or disadvantageous over the geometry shown in FIG. 1 and FIG. 13. All possible geometries that increase the surface area of the stationary and/or non-stationary walls of the compression chambers are envisioned and encompassed by this disclosure.

While not shown in FIG. 13, in some embodiments, a wobble plate compressor can further include a mechanism that causes the pistons to move or oscillate non-axially as the piston reciprocates axially, as is disclosed in reference to the compressor 100 and in other places herein. Any of the non-axial piston motions, and related mechanisms for causing such motions, described herein can also be incorporated in the modified versions of the compressor 400 or other wobble plate compressors.

As in other types of compressors, there are other ways (besides manipulating the stationary and/or non-stationary walls of the compression chambers) to increase the turbulence in the compression chambers of a wobble-plate compressor, in addition to increasing surface area within the compression chamber. Sculpting of the cooling surfaces in the chambers, for example, can increase turbulence locally. The opposing compression chamber surfaces can, for examples, be undulating, furrowed, dimpled, rough, polished, and/or have other characteristics.

As with other reciprocating compressors, any and all of the walls of the compression chambers of a wobble-plate compressor (either or both stationary and non-stationary) can serve as heat sinks to absorb and draw away the heat of compression. It is not necessary to use all three surface regions (the head, the cylinder side walls, and the piston) in this way, however. Significant improvement in efficiency can be gained by cooling only the stationary walls (the head and the side walls of the cylinder). In addition, cooling of both the stationary and non-stationary walls is also possible, and both can increase the efficiency of a wobble-plate compressor.

The head, the piston, and the cylinders side walls can be provided with significant externally applied cooling, to draw away the heat of the compression and to keep the compression chamber walls from heating up. The non-moving compression chamber walls (the head and cylinder side walls) can be air-cooled by cooling fins and optionally a fan on its exterior, as shown in FIG. 13. They can also be cooled by a cooling fluid, such as a water jacket.

In the example embodiment depicted in FIG. 13, the cooling of the compressor head and cylinders (and, indirectly, the cooling of the pistons) is accomplished through fan-cooled external cooling fins, as opposed to incorporation of a water jacket(s) or other means. The example external cooling fins are shown at the points labeled 424. The fan 426 draws in ambient air, or air that has been conducted from a cooler location, and blows it first across the radial external cooling fins that occupy the center portion of the head at the point labeled 428. The air then courses over a manifold including several inlet ports 430 and a similar manifold including several exhaust (outlet) ports 432. Such manifolding is not depicted in FIG. 13. The air then blows across the radial cooling fins on the outer portion of the head, at the points labeled 434, before turning approximately 90 degrees to travel in the path labeled 436 between the radially arrayed external cooling fins along case 438 (the sides) of the compressor.

Internally, in this example embodiment, internal cooling fins 440 are shown on the inside surface(s) of the case of the compressor. These, or otherwise-shaped, cooling surfaces afford greater surface area for absorbing heat from the oil (or other fluid), helping to keep the oil as close to ambient temperature as feasible. The heat absorbed from the oil is conducted to the outer surface of the compressor case and is partially conducted to the cooling surfaces (fins in the case of this example embodiment) thereon. FIG. 14 shows a possible arrangement of the external cooling fins 462 on the outside surface of the case 460 of the compressor and the internal cooling fins 464 on the inside surface of the case of the compressor. This exemplary arrangement is optional in the functioning of the compressor, but can increase the efficiency of removing heat from the case and the oil and component parts therein over other arrangements which do not align the internal and external cooling fins for improved heat conduction.

The non-stationary compression chamber walls (e.g., the pistons 412) can be cooled by being in occasional, continual, or continuous contact with a cooling fluid (e.g., liquid, vapor, or gas) of lower temperature, to which they can lose heat. The non-working sides of the pistons can be in contact with oil or other fluid that provides lubrication, in some embodiments. Such oil, and/or other fluids, can also be utilized as a heat sink, to extract significant heat from the pistons so that they can in turn serve in their role as heat sinks, to absorb and draw away some of the heat of compression that is produced in this, and all other types of gas compressors. FIG. 13 shows one possible approach as an example: spraying a cooling fluid 442 at the non-working sides of the pistons. In this example embodiment, use of spray nozzles or oil jets 444 to spray oil at the non-working sides of the pistons is shown. This use of fluid cooling of the non-stationary compression chamber walls of a wobble-plate compressor is another example of the general concept of cooling the non-working side of a moving wall of the compression chamber in order to use that wall to more efficiently extract heat from the compressing gas. FIG. 13 does not show either oil pump(s) or conducting tubes or other means to convey oil to the oil jets 444, though this can be accomplished in various ways.

As with the other reciprocating-type compressors discussed herein, the non-working sides of the non-stationary walls of the compression chambers (e.g., the pistons) can be shaped to increase the transfer of heat to the cooling fluid (e.g., liquid, vapor, or gas) to help to keep the pistons as close to ambient temperature as feasible. This shaping can, for example, include cooling fins, as shown in FIG. 13 at the points labeled 446. Many other geometries are possible for the non-working sides of the non-stationary walls of the compression chambers. Shaping of the non-working sides of the non-stationary walls of the compression chambers of a wobble-plate compressor to increase the surface area, or to otherwise increase the ability of the pistons to transfer heat to the cooling fluid (liquid, vapor, or gas), is another example of the concept of including increased surface area on the non-working side of a non-stationary wall of the compression chamber to increase heat conduction through and away from the non-stationary wall.

As shown in FIGS. 15A and 15B, gas can enter and exit the compression chambers through individual valve bodies 470 that function as both inlet and exhaust (outlet) valves 450 and 452. The valve 470 at configuration 450 is in the exhaust position, and the valve at 452 is in the intake position. The valves 470 can include spring mechanisms 472 that cause them the open and close in the proper functioning manner based on pressure differential on either side of the valve. The dimensions, including the relative proportions of the valve dimensions, as shown in FIG. 13, are exemplary for illustration and not necessarily optimized for efficient operation. The schematic representations of the valves shown in FIG. 13 depict the concept of concentric intake and exhaust (outlet) that can occupy one location as just one illustrative example.

FIGS. 16-18 illustrate embodiments of a scroll-type compressor 500 that include many of the features disclosed herein. The scroll-type compressor 500 can include two interlaid spiral plates (e.g., Archimedes spirals), with one of the spiral plates 502 serving as a stator and the other 504 serving as a rotor. Each spiral plate includes a flat disk portion with a spiral wall extending perpendicularly from one side of the disk portion, such that the spiral walls project towards each other and interleave, as shown in FIG. 16. The two disk portions are substantially parallel and form top and bottom walls of the compression chamber. The top and bottom disk portions are not shown in FIG. 16, only the spiral walls. The rotor 504 orbits about the center axis of the stationary stator 502. During operation, the two scrolls define a progression of compression chambers that form between the spirals of the stator and rotor, spiral inward, toward the centers of both the stator and the rotor, with decreasing volume, thus compressing the gas between them, and disappear at reaching the center of the compressor, at which point the compressed gas is expelled through an exhaust valve.

In FIGS. 17A and 17B, the basic functioning of a conventional scroll compressor is further illustrated. As the rotor 504 “orbits” the vertical center axis of the stator 502, without itself rotating, the vertical walls 507 of the rotor 504 oscillate back and forth between the vertical walls 506 of the stator 502. The moving vertical wall 507 can approach very close to both of the stationary walls 506 (e.g., within a few thousandths of an inch) but not actually touch the stationary walls 506. As in some other types of compressors (e.g., blowers, screw compressors, etc.), the non-stationary (i.e., moving) walls of the compression chamber need not actually contact the stationary walls of the compression chamber, though such contact is optional in some embodiments. This clearance between the stator 502 and rotor 504 is beneficial in a scroll-type compressor because, as the walls of the stator and rotor come into near proximity to one another, there is a sideways motion of the two relative to one another that can result in a scraping action if they touch.

However, in embodiments wherein the vertical walls 506, 507 of the stator and rotor do not touch, leakage can occur between them. Leakage between the vertical walls of the stator and rotor, at the interface labelled 508 (and other analogous interface in 18B) results in significant losses of efficiency. There is also leakage of the compressed and compressing gas at the interfaces labeled 510, where the tops of the rotor walls 507 interface with the bottom of the stator disk, and the bottoms of the stator walls 506 interface with the tops of the rotor disk, which interfaces extend over the entire spiral scroll length. At all of these interfaces, the compressed and compressing gas can leak out of the compression chambers toward lower pressure regions of the compressor.

The straight, vertical walls 506, 507 of the rotor and stator shown in FIGS. 17A and 17B are analogous to the flat piston and flat head of the conventional reciprocating piston-type compressor discussed herein. The flat vertical walls 506, 507 offer the minimum surface area to absorb heat from the compressed and compressing gas and avoid putting anything in the space reserved for the gas that is to be compressed.

As with reciprocating compressor embodiments disclosed herein, the walls of the compression chambers of a scroll-type compressor (either or both stationary and non-stationary) can also serve as heat sinks to absorb and draw away the heat of compression, heat that is produced in all compression of any compressible gas, even purely isothermal compression. It is not necessary to use both the stator and the rotor in this way. Significant improvement in efficiency can be gained by cooling those compression chamber walls that are easiest to cool, the stationary walls of the stator. But, cooling of both the stationary and non-stationary walls can further increase the efficiency of a scroll-type compressor.

FIGS. 18A and 18B shows an exemplary geometry for the scroll compressor 500, in which the purely vertical walls 506, 507 of the rotor and stator shown in FIGS. 17A and 17B are replaced with walls 512, 513 having much greater surface areas. As in the reciprocating compressors discussed herein, the vertical walls 512 of the stator 502 include projections extending perpendicularly from both sides of the wall, and the vertical walls 513 of the rotor 504 also includes projections extending perpendicular from both sides of the wall. Each opposing set of projections can be offset and correspondingly shaped such that the opposing projections can closely interleave with each other when the moving wall 513 approaches the stationary wall 512, similar to the reciprocating compressor embodiments and other compressor embodiments described herein. The illustrated projections of the walls 512, 513 can increase the surface area of these walls more that 400% compared to the flat faces of walls 506, 507. In alternative similar geometries, the surface can be increased by at least 200%, at least 300%, at least 400%, and/or at least 500%, compared to the flat faces of walls 506, 507. Many other geometries are possible at the interfaces between the moving and stationary vertical walls that increase the surface area of the walls of the compression chambers and thereby increase heat conduction between the gas in the chambers to the walls, and all such possible geometries are encompassed by this disclosure. In addition, the increased width of the vertical walls 512, 513 compared to the narrow width of the conventional straight vertical walls 506, 507, as seen at the points labeled 514, can substantially improve the ability of the vertical walls 512, 513 to transfer heat vertically to the disk portions of the stator and rotor that form the upper and lower horizontal walls of the compression chambers.

Many other geometries can alternatively be utilized to achieve the increased heat transfer away from the compressing gas. For example, some geometries can permit actual contact between the stator and rotor. Such contact between the rotor and stator offers the possibility of higher efficiency, by (1) reducing the losses of the compressed and compressing gas at the (near-contact) points of nearest approach, and (2) increasing turbulence within the compression chambers by improving the “squish” that happens at the points of near-contact or contact. Geometries that involve actual contact between the rotor and stator walls can include a mechanism that allows the rotor to rotate back-and-forth, or oscillate freely, a few degrees on its own axis, while orbiting the other scroll (the stator) about its axis. This motion can allow for the sideways (sliding) action at the points of near-contact (or of actual contact) that can be included in the orbiting motion of the one scroll (rotor) in the other scroll (stator). This manipulation of the rotor and/or stator, in ways not customary in the operation of a conventional scroll-type compressor, to allow actual contact between the two compression chamber walls, is another example of the concept described herein in various other examples, wherein the path of at least one compression chamber wall (rotor, stator, or both) has been altered from the conventional path to achieve increased gas flow and turbulence in the compression chamber, and thereby increased heat transfer within the gas and between the gas and the chamber walls.

The increased surface area of the example geometry shown in FIGS. 18A and 18B (and in many other alternative geometries for the cooling surfaces) and the shearing action that is created between these faces by the irregular sideways motion described above, creates significant turbulence of the gas that is caught between the rotor and stator and compressed as it moves toward the center of the spiral (scroll). As in other types of compressors described herein, there are additional and/or alternative ways to increase the turbulence in the compression chambers of a scroll-type compressor. Sculpting of the cooling surfaces, for example, can increase turbulence locally. Any of the surfaces of the compression chambers surfaces can, for examples, be undulating, furrowed, dimpled, rough, or polished.

In some embodiments, the stator or the rotor or both can be provided with substantial externally applied cooling, to draw away the heat of the compression and to keep the stator and/or rotor from heating up. As with a reciprocating compressor, the non-moving compression chamber walls (the stator) can be fluid-cooled, such as air-cooled by cooling fins and/or a fan on its exterior, or by a circulating cooling fluid, as in a water jacket. FIGS. 18A and 18B show an exemplary water jacket 516 located in the horizontal disk portion of the stator 502 to extract heat from the stationary walls of a scroll-type compressor.

As in other compressor embodiments disclosed herein, the non-stationary compression chamber walls (e.g., the rotor) can be cooled by being in occasional, continual, or continuous contact with a cooling fluid (liquid, vapor, or gas) of lower temperature, to which they can lose heat. Analogous to the configuration of many piston-type compressors, the non-working sides of rotors in many scroll-type compressors are in contact with oil that provides lubrication. However, in the disclosed scroll compressors, lubrication oil and/or other fluids serve as a substantial heat sink, to extract significant heat from the non-working side of the rotor so that it can in turn serve in its role as a heat sink, to absorb and draw away the heat of compression that is produced in the compression chambers. FIG. 18B shows one possible approach as an example: spraying a cooling fluid at the non-working side of the rotor from a nozzle 522. In some embodiments, such cooling fluid can be circulated through an external heat sink and returned cooled to be sprayed at the rotor again in a cyclical manner.

As with reciprocating-type compressors disclosed herein, the non-working side of the non-stationary walls of the compression chambers of a scroll-type compressor (e.g., the rotor) can be shaped to increase surface area and thereby increase the transfer of heat to the cooling fluid. This shaping can, for example, include cooling fins 518, as shown in FIGS. 18A and 18B, and/or other geometries. Shaping of the non-working side of the non-stationary walls of the compression chambers of a scroll-type compressor to increase the surface area, or to otherwise increase the ability of the rotor to transfer heat to the cooling fluid (liquid, vapor, or gas), is another example of the concept that can be applied analogously across many different species of compressors disclosed herein.

FIGS. 19-21 illustrate an exemplary low-pressure, high-flow, rotary blower-type compressor 600 (referred to herein simply as a blower), which incorporates many of the principles disclosed herein. Application of the disclosed principles in screw-type compressor embodiments (not illustrated) is analogous and encompassed by this disclosure. In the blower 600, two rotors 610 are counter-rotated by the shaft 612 (FIG. 19) of the compressor and two gears 614 (FIG. 19). In this manner, the rotors 610 capture a quantity of air at the inlet 616 (FIG. 20), and carry it along a path toward the outlet 618 (FIG. 20) of the blower, without substantial compression.

A blower is conventionally thought of as simply taking in air and conveying it to the outlet of the blower, and the compression that is achieved is conceptually thought of as happening outside of the compressor; the process can be described as involving “external compression,” as opposed to the “internal compression” that is characteristic of many other compressors. Even similar screw-type compressors can achieve compression within the compressor, and can be described as an “internal compression” type of compressor.

However, this conventional concept of a blower is not correct: the compression that is achieved by a blower occurs, though rather indirectly, within the blower. The compression occurs in what would otherwise be described as a compression chamber, even though the chamber does not compress the gas through operation of a decreasing compression chamber volume, as do many other types of compressors.

The chamber that captures a quantity of the gas to be compressed (and “blown”) can be seen at 620 in FIG. 20. In the chambers 620, a quantity of gas is carried along an angular path around the shaft 612, without compression. For example, in FIG. 20, the left-hand rotor 610 rotates counter-clockwise and the right-hand rotor rotates clockwise, such that the chambers 620 moves outward and down around the peripheral sides of the blower toward the outlet 618. At the point that a chamber 620 opens to the outlet, as shown at 624, it becomes a compression chamber, as it is connected to the outlet chamber and to the closing chamber of the opposing rotor, both of which are filled with higher pressure gas. The higher pressure is back-pressure, caused by a resistance downstream, that is, the load against which the blower is working. The higher pressure gas rushes into the “compression chamber” 620 when it joins at 624. In doing so, it does work against the gas that has been conveyed, in theory at essentially ambient pressure, to this point. The new gas is compressed as it rises to the pressure of the outlet; the outlet pressure falls momentarily, as previously compressed gas flows backward, into the “compression chamber” that has just opened. Thus, the compression actually happens inside the blower; it is not actually an “external compression” compressor.

Unlike other compressors that achieve compression of the gas by trapping it in a chamber of varying volume and compress the gas by reducing the volume of the compression chamber, in a blower, the equivalent of the non-stationary walls in other compressors, is a moving, higher-pressure gas, which effects the compression. Though it is not an accurate literal description, in a blower the “non-stationary walls” can be thought of as a pressure wave that compresses the newly admitted gas to a fraction of its original volume at 620. In reality, the newly admitted gas and the higher-pressure gas mix freely under the considerable turbulence that occurs when the chamber opens to the outlet. Like other conventional compressors, the compression-effecting feature in a conventional blower tends to be hotter than the new gas was when it was admitted to the inlet of the compressor at 616. Being hotter, the already compressed gas heats the newly admitted gas directly, while it is also being heated by the work of compression. This unfortunate feature of a conventional blower, as in other types of compressors, results in much of the inefficiency of a conventional blower.

The gas is also heated during the entire trip from the inlet 616 to the moment of opening of the chamber 620 to the outlet 618. This results from the general high temperature of the conventional blower, which heats up during operation, such as to temperatures higher than an internal combustion engine is allowed to reach. Thus, in a conventional blower, the gas is heated at essentially every moment between inlet and full compression.

If any of the direct heating of the gas in a blower can be avoided, that can improve the efficiency of the blower. If, in addition, the temperature rise that results from compressing the gas can be reduced by extraction of heat from the gas during the compression, efficiency can be further improved.

FIGS. 19-21 illustrate the application of the principles disclosed herein to a rotary-lobe blower. The blower shown has a straight-lobe design. The principles disclosed herein also apply equally to helical-lobe blowers.

The case of the blower can be fluid-cooled using a cooling jacket 626 in FIG. 20 (e.g., a water jacket). The end plates of the compressor can also be fluid-cooled, as shown at 630 in FIG. 19, such as by a portion of the cooling jacket that extends around the end plates. As in any application of the disclosed principles to any type of compressor, it can be valuable to cool any or all of the walls of the compressor, both the stationary and the non-stationary walls. In the blower 600, the non-stationary walls are the rotary lobes of rotors 610. In this embodiment, the rotary lobes can be hollow and can be cooled by flowing a cooling fluid through them, as shown at 632 in FIG. 19. The connection of the spinning shafts 612 to the coolant flow is accomplished through rotary unions 634. Within the lobes, the flow path of the coolant can be directed to be more general and to more effectively cool the surfaces of the lobes. In this embodiment, internal ducting of the coolant is shown at 640; while other means to direct the flow within the lobes can be included in other embodiments. A port 642 can be included to admit any gathered gas (air or other) within the lobes: a gas, being lower density than the cooling fluid, can migrate to the center and can escape into the shaft, from where it can exit the blower.

In the thoroughly cooled blower 600, rather than heating the gas during both its trip through the blower and during compression, the walls of the blower's chambers extract heat from the gas. The effectiveness of this cooling can be improved by addition of cooling fins, as shown in example at 644, or by other surface features that increase the surface areas of the lobes and the blower case. The example geometry shown in FIG. 21, which include close intermeshing between projections of the rotating lobes 660 and offset opposing projections of the stationary case 662, can increase the surface area of the lobes and the case by more than 700%. In other exemplary geometries, the increase in surface area at the lobe-case interface can be greater than 200%, greater than 300%, greater than 400%, greater than 500%, and/or greater than 600%. A limiting factor in adding surface area to the lobes and walls is that it may increase leakage of higher pressure gas backwards through the gap that is required between the lobes and the case. However, a sufficient seal between the lobes and the blower case can be maintained by configuring the intermeshing cooling surfaces to fit closely, as shown in FIG. 21.

As noted herein, it can be valuable to cool all of the “walls” of the blower, so that the gas is not heated by the blower, or such heating is minimized. As also noted herein, the compression-effecting feature of a blower is the high-pressure gas in the outlet rushing into the chamber on opening. As in other types of compressors, it is valuable to effectively cool the compressing feature of a blower. This is analogous to cooling the piston in a reciprocating compressor, or cooling the scrolls and plates of a scroll-type compressor. To effect the sufficient cooling of the higher-pressure gas in a blower, a heat sink 650 (FIG. 20) can be positioned in the outlet 618 of the blower. A large rectangular or similarly shaped outlet port can be included here, rather than a circular or otherwise smaller outlet port used in some conventional blowers; this can allow the heat sink 650 to serve its cooling function along the entire length, or most of the length, of the rotors.

The heat sink 650 may be in the form of a radiator, as shown, and/or in the form of plates that are integral with the blower case or thermally connected to the blower case, which can then be cooled by an external heat sink, such as a water jacket or cooling fins. In some embodiments, the heat sink, whether integral or not, can be cooled by flowing the fluid in the cooling jacket through the heat sink. A bar and plate type radiator shown at 650 may instead be a tube and fin type radiator. The plates or tubes of a radiator, if any, can also be placed transversely, instead of longitudinally, as shown; this can also be efficient in cooling the higher pressure gas as it enters the alternately opening chambers, in allowing the left-right component of the flow as the chambers of the two rotors open alternately, and in facilitating use of the cooling jacket fluid to cool the heat sink. In any configuration, the location of the heat sink in the outlet can be valuable in removing heat from the gas and increasing efficiency.

It can be desirable for the heat sink be located close to the faces of the spinning lobes; that is, the heat sink may be more effective when closely coupled to the lobes. Placement of radiators, sometimes known as after-coolers, downstream of a blower cannot serve the function described herein in cooling the gas that is about to back up into the opening chambers. Because the volume of gas that flows backward into the opening chamber is less than or not much greater than the volume of the chamber, the cooling that is required must be accomplished very close to the opening point of the chamber, so that all or nearly all of the gas that enters the chamber has been cooled. The higher pressure gas in the outlet that is going to compress the gas in the chamber, on opening, is desirably cooled before it enters the chamber, as is the case in the blower 600.

Many of the features and principles disclosed herein can be applied to and included in various other embodiments that those to which they refer in this disclosure, and can be so applied or included in any practical combination, all of which embodiments are expressly included and encompassed by this disclosure. Such features and principles include, in addition to others disclosed elsewhere herein, the following:

(a) Concentric valves, such as those shown in the Figures depicting reciprocating compressors and wobble-plate compressors.

(b) Operating compressors at faster rates, such as above 1000 rpm or 2000 rpm. Conventional compressors run fairly slowly, at least compared to gasoline and diesel engines, such as about 900 rpm, just above an idle for most engines. The main reason is that the temperature rise of the compressor would be even worse, and most compressors could not be run continuously. This is not a problem solely for reciprocating compressors. Scroll compressors can also include excessive temperature rise as a particular problem that keeps them from being run faster. Without the cooling of the non-stationary surfaces, compressor cannot be operated very fast, even if the stationary surfaces are fluid-cooled. Thoroughly cooled as they are, the disclosed compressors can be run faster than conventional compressors.

(c) Operating compressors at higher pressures than is possible with conventional compressors, due to the ability to maintain the compressor at that temperature without the temperature increasing further.

(d) Compressors that take advantage of additional, even what might be called excessive, cooling. If there is a source of coldness available, such as water from a well, or water cooled in a well, or, in wintertime a radiator placed outside, a compressor can be cooled to below ambient air temperature, taking advantage of the low temperature source. Without the features and principles disclosed herein, even the extra coldness of the outdoor air cannot be used to its full advantage. With the disclosed technology, it is possible to cool a compressor down, such as below the ambient air temperature, and convey that lower temperature into the compression chamber and this to the compressing gas, to the point that the work of compression falls to exceptionally low levels.

(e) Cooling the scrolls of a scroll compressor. For refrigerators, at least, conventional compressors are sealed up in a can and get no cooling whatever. Cooling the scrolls provides a huge efficiency advantage.

(f) Cooling fins on the non-working sides of the non-stationary scroll walls.

(g) For blowers: positioning a radiator directly in the outlet port very close to the lobes, as noted elsewhere herein, instead of farther down the outlet conduit like conventional after-coolers.

(h) For blowers and other rotary compressors (e.g., screw compressors): fluid-cooling of the rotors and the axial ends of the blower cases.

(i) Continuously cooling the non-stationary walls of any compressor.

Conventional compressors have not been able to take significant advantage of the cooling that has been applied to them, due to:

(a) lack of sufficient surface area in the compression chamber walls to which the compressing fluid can lose heat during the compression;

(b) the walls of the compression chamber, especially the non-stationary walls of the compression chamber, not being cooled sufficiently to serve in the role of extracting heat from the fluid that is being compressed, primarily during the compression; and

(c) lack of sufficient turbulence in the compressing fluid to bring it into effective thermal contact with the walls of the compression chamber, such that the fluid can lose heat to the walls of the compression chamber, during the compression, at the rates required to effect near-isothermal compression of the fluid.

The technology disclosed herein solves all of these problems (though not every disclosed embodiment necessarily solves all of these problems) and, for the first time, results in compressors that can take fuller advantage of external cooling, specifically to change the compression process over the short time scales involved, and thereby effect compression that is significantly closer to isothermal compression and significantly more efficient in the use of energy. For the first time, compressors can make effective use of external cooling, even cooling below ambient temperature when reasonably possible, to change the nature and progression of the compression process itself, moving it to near-isothermal compression, and possibly, with sufficient cooling, even beyond the theoretical efficiency of isothermal compression.

Additional exemplary claims include:

A. A reciprocating piston-type compressor, comprising:

a housing;

a piston that reciprocates within the housing along an axial dimension of the compressor, the housing and the piston having respective wall surfaces which together form a compression chamber that varies in volume based on the axial position of the piston relative to the housing;

wherein the housing includes a fluid inlet for receiving into the compression chamber a fluid to be compressed and a fluid outlet for expelling out of the compression chamber the fluid in a compressed state;

wherein the housing comprises first projections that project into the compression chamber;

wherein the piston comprises a piston head having a working side facing the compression chamber and a non-working side facing away from the compression chamber, wherein the working side of the piston head includes second projections that project into the compression chamber opposite the first projections;

wherein the second projections are oriented such that they interleave between the first projections as the piston reciprocates, whereby the first and second projections increase the surface area of the wall surfaces and increase heat conduction away from the fluid in the compression chamber when compared to a cylindrical compression chamber; and

wherein the first projections and the second projections are each arrayed circumferentially around a common central axis, the first and second projections extend both axially and radially away from the central axis, and the first and second projections increase in circumferential thickness with increasing radial distance from the central axis.

B. The compressor of claim A, wherein the first and second projections have a generally trapezoidal-shaped cross-sectional shape when viewed in a plane perpendicular to the axial direction.

C. The compressor of claim A, wherein the second projections have the same shape and size as the first projections, the second projections fit flushly in corresponding valleys between the first projections, and the second projections are offset circumferentially relative to the first projections about one half the angular width of one of the first projections.

D. The compressor of claim A, wherein the housing has a cylindrical side wall and the first projections are integral with the cylindrical side wall such that heat is conducted radially through the first projections into the sidewall.

In view of the many possible embodiments to which the principles of the disclosed technology may be applied, it should be recognized that the illustrated embodiments are only examples and should not be taken as limiting the scope of the disclosure. Rather, the scope of the disclosure is at least as broad as the following claims. I therefore claim all that comes within the scope of these claims. 

1. A fluid compressor, which comprises: a housing; one or more compression chambers that comprise both stationary and non-stationary walls that define the compression chambers; wherein the walls of the compression chambers are formed and structured to have at least twice the surface area that the compression chamber walls of a conventional compressor of the type has, and the stationary and non-stationary walls of the compression chambers partially or perfectly interleave or nest, or both, such that the volume of the compression chambers can be reduced sufficiently to effect a desired level of compression; wherein the walls of the compression chambers are formed and structured such that a larger portion of the fluid, compared to a conventional compressor of the type, is close to one or more walls of the compression chambers for most of the time during which the fluid is being compressed, such that said portion of the fluid is in better thermal contact with one or more walls of the compression chamber than in a conventional compressor of the type; wherein at least some of the stationary compression chamber walls are cooled by extraction of heat from said walls by, or to, one or more external heat sinks, such that said walls remain near or below the ambient temperature of the fluid to be compressed prior to entering the compressor; wherein at least some of the non-stationary compression chamber walls are cooled by extraction of heat from said non-stationary walls by fluid cooling, the extracted heat being conveyed by the cooling fluid to one or more heat sinks, which heat sinks may be external to the compressor, such that at least some of the non-stationary walls of the compressor remain near or below the ambient temperature of the fluid prior to entering the compressor; one or more compressing features that reduce the volume of the fluid that is to be compressed and thereby effect the compression of said fluid, which compressing features may be the non-stationary walls of the compression chambers, which non-stationary walls reduce the volume of the compression chambers through their motion and thereby effect the compression of the fluid wherein the one or more compressing features are cooled by extraction of heat from said compressing features by, or to, one or more heat sinks, such that said compressing features, remain near or below the ambient temperature of the fluid to be compressed prior to entering the compressor.
 2. The compressor of claim 1, wherein the manner or path in which the non-stationary walls of the compression chambers move or are moved is different from the manner or path of non-stationary walls in a conventional compressor of the type; wherein said different manner or path of movement of the non-stationary walls of the compression chambers results directly or indirectly in increased intra-chamber flow of the fluid to be compressed or results in increased turbulence in the fluid to be compressed.
 3. The compressor of claim 1, wherein the manner and/or path in which the non-stationary walls of the compression chambers move or are moved is that manner and path that are typical of a wobble-piston compressor; wherein said wobble piston manner and path of movement of the non-stationary walls of the compression chambers results directly or indirectly in increased intra-chamber flow of the fluid to be compressed and/or results in increased turbulence in the fluid to be compressed.
 4. The compressor of claim 2, wherein the compressor is a reciprocating piston type compressor; wherein the stationary and non-stationary walls of the compression chamber or chambers have a geometry that is substantially rotationally symmetric, and the different manner or path of movement of the non-stationary walls of the compression chambers includes rotationally oscillatory motion, and results in increased intra-chamber flow of the fluid being compressed or results in increased turbulence in the fluid being compressed.
 5. The compressor of claim 1, wherein fluid cooling is not provided for the non-stationary walls of the compression chambers.
 6. A reciprocating piston-type compressor, comprising: a housing; a piston that reciprocates within the housing along an axial dimension of the compressor, the housing and the piston having respective wall surfaces which together form a compression chamber that varies in volume based on the axial position of the piston relative to the housing; wherein the housing includes a fluid inlet for receiving into the compression chamber a fluid to be compressed and a fluid outlet for expelling out of the compression chamber the fluid in a compressed state; wherein the housing comprises first projections that project into the compression chamber; wherein the piston comprises a piston head having a working side facing the compression chamber and a non-working side facing away from the compression chamber, wherein the working side of the piston head includes second projections that project into the compression chamber; wherein the second projections are oriented such that they interleave between the first projections as the piston reciprocates, and the first and second projections increase the surface area of the wall surfaces and increase heat conduction away from the fluid in the compression chamber when compared to a cylindrical compression chamber; and wherein the compressor further comprises a cooling fluid that is conducted into contact with the non-working side of the piston head to conduct heat away from the piston head via the cooling fluid.
 7. The compressor of claim 6, wherein the cooling fluid is sprayed against the non-working side of piston head.
 8. The compressor of claim 6, wherein the non-working side of piston head includes cooling fins that increase the surface area of the non-working side of the piston and the cooling fluid is conducted into contact with the cooling fins to conduct heat away from the piston.
 9. The compressor of claim 6, further comprising a cooling fluid jacket positioned around the housing to conduct heat away from the housing.
 10. The compressor of claim 6, wherein the first and second projections are arrayed circumferentially around a common central axis and the first and second projections increase in circumferential thickness as a function of increasing radial distance from the central axis.
 11. The compressor of claim 6, wherein the second projections oscillate in a non-axial direction relative to the first projections while the piston reciprocates axially within the housing, thereby increasing turbulence within the fluid and further increasing heat conduction away from the fluid.
 12. The compressor of claim 6, wherein the first projections are integral with side walls of the housing.
 13. The compressor of claim 1, wherein the fluid enters and exits the compression chamber via a common central port opposite from the piston.
 14. A reciprocating piston-type compressor, comprising: a housing; a piston that reciprocates within the housing along an axial dimension of the compressor, the housing and the piston having respective wall surfaces which together form a compression chamber that varies in volume based on the axial position of the piston relative to the housing; wherein the housing includes a fluid inlet for receiving into the compression chamber a fluid to be compressed and a fluid outlet for expelling out of the compression chamber the fluid in a compressed state; wherein the housing comprises first projections that project into the compression chamber; wherein the piston comprises a piston head having a working side facing the compression chamber and a non-working side facing away from the compression chamber, and the working side of the piston head includes second projections that project into the compression chamber opposite the first projections; wherein the second projections are oriented such that they interleave between the first projections as the piston reciprocates, and the first and second projections increase the surface area of the wall surfaces and increase heat conduction away from the fluid in the compression chamber when compared to a cylindrical compression chamber; and wherein the piston and the second projections oscillate in a non-axial direction relative to the housing and the first projections while the piston reciprocates axially within the housing, thereby increasing turbulence within the fluid and further increasing heat conduction away from the fluid.
 15. The compressor of claim 14, wherein the piston moves through at least one full oscillation during each axial compression stroke of the piston.
 16. The compressor of claim 14, wherein the oscillation of the piston is caused by a mechanical interface between the piston and the housing.
 17. The compressor of claim 14, wherein the oscillation of the piston and second projections comprises rotational motion of the piston about a central axis of the compressor.
 18. The compressor of claim 14, wherein the first and second projections are arrayed circumferentially around a common central axis and the first and second projections increase in circumferential thickness as a function of increasing radial distance from the central axis.
 19. The compressor of claim 14, wherein each of the second projections oscillates in a corresponding valley between two of the first projections during the axial reciprocation of the piston.
 20. The compressor of claim 14, wherein the first and second projections have a generally trapezoidal cross-sectional shape when viewed in a plane perpendicular to the axial direction. 